Control valve for variable displacement compressor

ABSTRACT

In a control valve for a variable displacement compressor, for controlling the discharge flow rate of refrigerant to be constant, to dispense with a check valve in a refrigerant outlet port of the compressor. The control valve includes a first control valve that controls the passage cross-sectional area of a refrigerant passage through which refrigerant passes from a discharge chamber of the compressor to a refrigerant outlet port thereof, and a second control valve that controls the flow rate of refrigerant allowed to flow from the discharge chamber to a crankcase such that a differential pressure (Pdh−Pdl) across the first control valve generated by the refrigerant passing therethrough becomes constant. The control valve is configured such that when a solenoid section is not energized, a first valve element is engaged with a piston to forcibly fully open the second control valve. The piston has an outer diameter equal to the inner diameter of a second valve seat, so that the discharge pressure Pdl on a refrigerant outlet port side is inhibited from adversely affecting the fully-opening operation of the second control valve, to thereby maintain the fully-closed state of the first control valve. This makes it possible to dispense with a check valve conventionally provided in the refrigerant outlet port.

CROSS-REFERENCE TO RELATED APPLICATION, IF ANY

This application claims priority of Japanese Application No. 2006-004395 filed on Jan. 12, 2006, entitled “Control Valve For Variable Displacement Compressor”, No. 2006-039365 filed on Feb. 16, 2006, entitled “Control Valve For Variable Displacement Compressor”, and No. 2006-238904 filed on Sep. 4, 2006, entitled “Control Valve For Variable Displacement Compressor”.

BACKGROUND OF THE INVENTION

(1) Field of the Invention

The present invention relates to a control valve for a variable displacement compressor, and more particularly to a control valve for a variable displacement compressor, for controlling a flow of refrigerant discharged from the compressor to a constant flow rate.

(2) Description of the Related Art

As a compressor used in the refrigeration cycle of an automotive air conditioner, for compressing refrigerant, a variable displacement compressor capable of varying the volume (discharge amount) of refrigerant is employed so as to obtain an adequate cooling capacity without being constrained by the rotational speed of an engine which is the drive source of the compressor. In such a variable displacement compressor, pistons that reciprocate in parallel with a rotational shaft driven by the engine for rotation are connected to a wobble plate (swash plate) fitted on the rotational shaft, and by rotating the wobble plate while varying the inclination angle thereof within a crankcase, the stroke of the pistons is varied to control the capacity of the compressor, that is, the discharge amount of refrigerant.

In order to change the inclination angle of the wobble plate, the balance of pressures acting on the both sides of each piston connected to the wobble plate is changed by introducing part of compressed refrigerant into a hermetically closed crankcase to cause a change in the pressure in the crankcase.

In general, pressure in the crankcase is changed by controlling a control valve for the compressor provided in a passage communicating between the discharge chamber and the crankcase such that communication through the control valve is allowed or blocked. Now, when the control valve is set to a predetermined valve lift, if the rotational speed of the engine increases, pressure introduced from the discharge chamber into the crankcase increases to make the inclination angle of the wobble plate close to an angle perpendicular to the rotational shaft, whereby the volume of compressible refrigerant is controlled to be small. Inversely, when the rotational speed of the engine lowers, pressure introduced into the crankcase decreases whereby the volume of compressible refrigerant is controlled to be large. Thus, the variable displacement compressor is controlled such that the volume of discharged refrigerant is not varied irrespective of the rotational speed of the engine.

As methods for controlling the displacement of a variable displacement compressor using a control valve therefor, it is generally known, for example, to cause suction pressure Ps in the suction chamber to be held constant, and to cause the differential pressure between the suction pressure Ps in the suction chamber and a discharge pressure Pd in the discharge chamber to be held constant. It is also known to cause the flow rate of refrigerant discharged from the compressor to become constant (see e.g. Japanese Unexamined Patent Publication No. 2001-107854 (Paragraph Nos. [0035] to [0036], and FIG. 3)).

According to this control valve for a variable displacement compressor, disclosed in Japanese Unexamined Patent Publication No. 2001-107854, the differential pressure between two pressure monitoring points is detected by sensors to thereby indirectly grasp the flow rate of refrigerant drawn into the suction chamber, and the control valve controls the flow rate of refrigerant introduced from the discharge chamber into the crankcase such that the flow rate of refrigerant drawn into the suction chamber becomes constant, whereby the flow rate of refrigerant discharged from the compressor is controlled to be constant.

In contrast, a control valve for a variable displacement compressor is also known which dispenses with the sensors for detecting the differential pressure between two pressure monitoring points (see e.g. Japanese Unexamined Patent Publication No. 2004-116349 (Paragraph Nos. [0102] to [0108], and FIG. 12)). This control valve comprises a first control valve that controls the flow rate of refrigerant allowed to flow from a discharge chamber of the compressor to a refrigerant outlet port of the compressor, a second control valve that senses the differential pressure across the first control valve using a diaphragm thereof and controls the flow rate of refrigerant allowed to flow from the discharge chamber into a crankcase of the compressor based on the differential pressure, to change the displacement of the compressor, to thereby control the flow rate of refrigerant which the first control valve allows to flow to be constant, and a solenoid section that sets the flow rate of refrigerant to be allowed to flow by the first control valve, all of which are arranged along the same axis. According to this control valve, the first control valve forms a variable orifice that has its passage area of a refrigerant passage set by the solenoid section according to changes in external conditions, and the second control valve senses the differential pressure across the variable orifice, and controls pressure in the crankcase such that the differential pressure becomes equal to a predetermined value. As a result, the differential pressure across the variable orifice set to a certain passage area is held at the predetermined value, whereby the flow rate of refrigerant discharged from the compressor is controlled to be constant.

In the conventional control valve disclosed in the above-described Japanese Unexamined Patent Publication No. 2004-116349, when the compressor stops its operation, its capability of compressing and discharging refrigerant is suddenly lost to invert the relationship in pressure between the discharge chamber that has been at high pressure and the refrigerant outlet port located downstream of the first control valve. This acts not to control the second control valve to the minimum displacement side but to control the same to the maximum displacement side. To solve this problem, the conventional control valve is configured assuming that a check valve is provided at the refrigerant outlet port so as to prevent the first control vale from being adversely affected by the pressure at the refrigerant outlet port upon stoppage of the compressor. Therefore, the variable displacement compressor using the conventional control valve suffers from the problem of increased manufacturing costs due to the necessity of provision of the check valve.

SUMMARY OF THE INVENTION

The present invention has been made in view of the above problems, and an object thereof is to provide a control valve for a variable displacement compressor, which is of a type that controls the flow rate of discharged refrigerant, without requiring a check valve to be provided at a refrigerant outlet port of the compressor.

To solve the above problem, the present invention provides a control valve for a variable displacement compressor, including a first control valve that controls a flow rate of refrigerant which the first control valve allows to flow from a discharge chamber of the compressor to a refrigerant outlet port of the compressor, a second control valve that controls a flow rate of refrigerant which the second control valve allows to flow from the discharge chamber into a crankcase of the compressor based on a differential pressure across the first control valve, to change a displacement of the compressor, to thereby control the flow rate of the refrigerant allowed to flow by the first control valve to be constant, and a solenoid section that sets the flow rate of refrigerant which the first control valve is to allow to flow, wherein the control valve is made insensitive to pressure on a downstream side of the first control valve, and when the solenoid section is in a non-energized state, the first control valve is in a fully closed state, and the second control valve is in a fully open state, the second control valve is forcibly held in the fully open state even when the pressure on the downstream side of the first control valve is equal to or higher than pressure on an upstream side of the first control valve.

The above and other objects, features and advantages of the present invention will become apparent from the following description when taken in conjunction with the accompanying drawings which illustrate preferred embodiments of the present invention by way of example.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a central longitudinal cross-sectional view of the whole construction of a control valve for a variable displacement compressor, according to a first embodiment.

FIG. 2 is a partial enlarged cross-sectional view of the construction of a valve section of the control valve according to the first embodiment.

FIG. 3 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a second embodiment.

FIG. 4 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a third embodiment.

FIG. 5 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a fourth embodiment.

FIG. 6 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a fifth embodiment.

FIG. 7 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a sixth embodiment.

FIG. 8 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a seventh embodiment.

FIG. 9 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to an eighth embodiment.

FIGS. 10A to 10C are views for explaining characteristics of a diaphragm, in which FIG. 10A shows a state in which no differential pressure is applied to the diaphragm, FIG. 10B shows a state in which the diaphragm is displaced by a differential pressure, and FIG. 10C shows a state in which the differential pressure is applied to the displaced diaphragm in a direction opposite to the direction of the displacement.

FIGS. 11A to 11C are explanatory views showing the construction of a differential pressure-sensing section of the control valve according to the eighth embodiment, in which FIG. 11A shows a case where pressure in a discharge chamber of the compressor is higher than pressure in a refrigerant outlet port of the compressor, and FIG. 11B shows a case where the pressure in the discharge chamber is lower than the pressure in the refrigerant outlet port.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

FIG. 1 is a central longitudinal cross-sectional view of the whole construction of a control valve for a variable displacement compressor, according to a first embodiment of the present invention. FIG. 2 is a partial enlarged cross-sectional view showing the construction of a valve section of the control valve according to the first embodiment.

The control valve 10 for a variable displacement compressor comprises a first control valve 10A that controls a passage cross-sectional area of a refrigerant passage through which high-pressure refrigerant flows from a discharge chamber of the compressor to a refrigerant outlet port thereof, a second control valve 10B that controls the flow rate of refrigerant to be supplied from the discharge chamber to a crankcase of the compressor, and a solenoid section 10C that sets the passage cross-sectional area of the refrigerant passage of the first control valve 10A, all of which are arranged on the same axis.

The first control valve 10A and the second control valve 10B have a first body 11 and a second body 12 press-fitted into the first body 11. The first body 11 and the second body 12 are provided with a port 13, a port 14, and a port 15. When the control valve 10 is mounted in the compressor, the port 13, the port 14, and the port 15 are communicated respectively with the discharge chamber of the compressor, for introducing refrigerant at discharge pressure Pdh into the first control valve 10A, with the refrigerant outlet port of the compressor, for discharging refrigerant at discharge pressure Pdl from the first control valve 10A, and with the discharge chamber of the compressor, for introducing refrigerant at discharge pressure Pdh2 into the second control valve 10B. The second body 12 has foremost end formed with a port 16 that is communicated with the crankcase of the compressor, for discharging refrigerant at pressure Pc from the second control valve 10B.

Although the control valve 10 can be applied to a variable displacement compressor configured such that when the control valve 10 is mounted therein, the port 13 at the discharge pressure Pdh and the port 15 at the discharge pressure Pdh2 both communicate with the discharge chamber thereof, it is preferable that the control valve 10 is applied to a variable displacement compressor configured such that the port 13 at the discharge pressure Pdh is directly communicated with the discharge chamber, and the port 15 at the discharge pressure Pdh2 is communicated with an outlet port of an oil separator disposed on the downstream side of the discharge chamber. This enables the second control valve 10B to return compressor lubricating oil, contained in refrigerant in a large amount, while controlling pressure Pc in the crankcase.

The first control valve 10A has a passage axially formed through the second body 12 such that the passage communicates between the port 13 at the discharge pressure Pdh and the port 14 at the discharge pressure Pdl. A first valve seat 17 is rigidly press-fitted in the passage, and a first valve element 18 is disposed on the downstream side of the first valve seat 17 in a manner movable to and away from the first valve seat 17.

The first valve element 18 has a hollow cylindrical portion integrally formed in a manner axially extending through a valve hole, and a guide 19 is rigidly press-fitted in the hollow cylindrical portion. The guide 19 is urged by a spring 20 in the valve-closing direction of the first control valve 10A. The guide 19 is configured such that a portion thereof in sliding contact with an inner wall of the first body 11 has an outer diameter equal to the inner diameter of the first valve seat 17, whereby the discharge pressure Pdh introduced into the port 13 equally acts on the first valve element 18 and the guide 19 in respective opposite directions to prevent the discharge pressure Pdh from adversely affecting the control operation of the first control valve 10A.

The guide 19 has a refrigerant passage axially extending therethrough, and is provided with a check valve 21 for opening and closing the refrigerant passage. The check valve 21 has a valve element 22, e.g. made of rubber, disposed on a low-pressure side of the refrigerant passage formed in the guide 19, that is, on the side communicating with the port 14 at the discharge pressure Pdl via the first valve element 18, and a leaf spring 23 for axially movably holding the valve element 22. In a neutral state of the valve element 22 in which no pressure acts thereon, the valve element 22 is held at a position where the refrigerant passage is slightly open, by the leaf spring 23.

The second control valve 10B has a second valve seat 31 that is press-fitted into the foremost end of the second body 12, which has the port 16 formed therethrough in the axial direction, and a second valve element 32 disposed on the upstream side of the second valve seat 31 in a manner movable to and away from the second valve seat 31 such that the second valve element 32 opens and closes the port 16 on the upstream side. The second valve element 32 is axially movably held by the second body 12, and is mounted on the piston 33 forming a differential pressure-sensing section. More specifically, the piston 33 is configured such that it has a valve element base portion-accommodating portion 34 recessed in a surface thereof opposed to the second valve seat 31, for having a spring 35 and a base portion of the second valve element 32 arranged in the accommodating portion 34, and an open end of the accommodating portion 34 is swaged to prevent the second valve element 32 from being pushed out by the spring 35. This makes it possible to soften impact of collision which is caused between the second valve element 32 and the second valve seat 31 when the second control valve 10B is fully closed quickly.

Further, the piston 33 is urged in the valve-closing direction of the second control valve 10B by a spring 37 disposed between the piston 33 and a spring-receiving portion 36 rigidly press-fitted into the second body 12 from the side toward the port 14. The spring 37 is set to a spring force smaller than that of the spring 20 urging the first control valve 10A in the valve-closing direction. Furthermore, the piston 33 is integrally formed with an extended portion 38 extended therefrom toward the solenoid section 10 c into the first valve element 18. A washer 39 is fixed to an end of the extended portion 38 by swaging. The washer 39 is brought into engagement with a stepped portion formed on the first valve element 18. Thus, when the first control valve 10A is fully closed, the piston 33 is forcibly pulled by the first valve element 18 in the valve-opening direction of the second control valve 10B, and hence the second control valve 10B can be held in the fully open state. Further, the piston 33 is configured to have an outer diameter equal to the inner diameter of the first valve seat 17 such that when the piston 33 is engaged with the first valve element 18, the piston 33 is prevented from being adversely affected by the discharge pressure Pdl.

Furthermore, the second control valve 10B has film-like seal rings 40 and 41 formed e.g. of rubber for sealing clearances between the piston 33 and the second body 12 by pressures in the ports 14 and 15. The seal rings 40 and 41 are arranged between a stepped portion of the second body 12 and the spring-receiving portion 36 and between a stepped portion of the second body 12 and the second valve seat 31, respectively.

The solenoid section 10 c has a core 51 rigidly press-fitted into a central opening of the first body 11. The core 51 is fitted into an opening of a bottomed sleeve 52 in a manner blocking the opening. The bottomed sleeve 52 contains a plunger 53, a shaft 54 axially extending through the core 51 and rigidly fixed to the plunger 53, an adjustment member 55 disposed on a bottom of the bottomed sleeve 52 for axially plastically deforming the bottom, thereby adjusting spring loads, a spring 56 disposed between the core 51 and the plunger 53, and a spring 57 disposed between the plunger 53 and the adjustment member 55. The shaft 54 is axially movably held by the core 51 and the plunger 53, with a free end thereof extending into the guide 19. When the solenoid section 10 c is energized, the free end is brought into abutment with an intercommunicating plate 24 fitted on a side of the guide 19 opposite from a side where the check valve 21 is disposed for the refrigerant passage formed in the guide 19, for urging the first valve element 18 in the valve-opening direction of the first control valve 10A. Arranged around the outer periphery of the bottomed sleeve 52 are a coil 58 and a yoke 59.

On the outer peripheries of the first body 11 and the second body 12, there are circumferentially provided an O ring 61 for sealing between the port 13 and the port 14, an O ring 62 for sealing between the port 14 and the port 15, an O ring 63 for sealing between the port 15 and the port 16, and an O ring 64 for sealing between the port 13 and the atmosphere, when the control valve 10 is mounted in the compressor.

In the control valve 10 configured as above, when the compressor is driven for rotation by the driving force of the engine, the compressor draws refrigerant from a suction chamber for compression, and discharges the compressed refrigerant.

At this time, when the solenoid section 10C is not energized, as shown in FIG. 1, the first control valve 10A is forcibly fully closed by the urging force of the spring 20, and the second control valve 10B is fully open since the piston 33 is pulled in the valve-opening direction against the urging force of the spring 37 by the first valve element 18. Therefore, since all the refrigerant discharged from the discharge chamber is introduced into the crankcase via the second control valve 10B, the compressor is in the minimum displacement operation state. As described above, when the solenoid section 10C is not energized, the compressor is in the minimum displacement operation state, so that the control valve 10 can be applied to a variable displacement compressor which does not necessitate an electromagnetic clutch for performing On-Off control of transmission of the driving force between the compressor and the engine for driving the compressor for rotation.

Now, when the compressor is started, control current is supplied to the solenoid section 10C. As the control current increases, the plunger 53 is pulled by the core 51, whereby the first valve element 18 is pushed upward by the shaft 54, as viewed in FIG. 1. In accordance with the upward motion of the first valve element 18, the piston 33 of the second control valve 10B, engaged with the first valve element 18, is also pushed upward, as viewed in FIG. 1, by the spring 37, until the second valve element 32 is seated on the second valve seat 31 to fully close the second control valve 10B. As a result, since all the refrigerant discharged from the discharge chamber ceases to be introduced into the crankcase, the compressor is now shifting to the maximum displacement operation.

When the control current is further increased, the first valve element 18 continues to be lifted, but in the second control valve 10B, which has been shifting in the valve-closing direction along with the lift of the first valve element 18, the motion of the piston 33 is stopped by the second valve element 32 being seated on the second valve seat 31, and therefore the first valve element 18 is disengaged from the piston 33.

After that, when the control current is held at a predetermined value, the first valve element 18 is stopped at a position where the urging force of the solenoid section 10C corresponding to the predetermined value and the urging force of the spring 20 against the solenoid force are balanced. The position where the first valve element 18 is stopped does not change until the value of the control current is changed. As described above, the first valve element 18 is stopped after being lifted from the first valve seat 17, whereby the first control valve 10A is set to a predetermined passage cross-sectional area with respect to the refrigerant passage thereof, to allow refrigerant at the discharge pressure Pdh, introduced into the port 13, to flow through the refrigerant passage having the predetermined passage cross-sectional area, so that refrigerant at the discharge pressure Pdl is discharged from the port 14.

When refrigerant flows through the first control valve 10A, a predetermined differential pressure (Pdh−Pdl=ΔP) is generated across the first control valve 10A. Since the discharge pressure Pdh2 of refrigerant supplied to the port 15 of the second control valve 10B is approximately equal to the discharge pressure Pdh of refrigerant supplied to the port 13 of the first control valve 10A, the differential pressure ΔP generated across the first control valve 10A can be sensed by the piston 33.

Up to this time point, the compressor has been operating toward its maximum displacement operation, so that the discharge pressure Pdh presently increases, and the flow rate of refrigerant passing through the first control valve 10A also increases. When the flow rate of refrigerant becomes larger than a predetermined value, an urging force in the valve-opening direction, caused by application of the differential pressure (Pdh2−Pdl≈ΔP) to the piston 33, comes to overcome the urging force of the spring 37, causing the second valve element 32 to be moved away from the second valve seat 31 to open the second control valve 10B. This causes refrigerant discharged from the discharge chamber to be introduced into the crankcase, which starts the variable displacement of the compressor.

After that, when the flow rate of refrigerant passing through the first control valve 10A increases to increase the differential pressure A P across the first control valve 10A, the piston 33 senses a change in the differential pressure to further open the second control valve 10B, and controls the compressor in a direction of decreasing the displacement thereof. Further, when the flow rate of refrigerant flowing through the first control valve 10A decreases to decrease the differential pressure ΔP across the first control valve 10A, the piston 33 senses a change in the differential pressure to cause the second control valve 10B to shift in the valve-closing direction, and controls the compressor in a direction of increasing the displacement thereof.

At this time, although in the second control valve 10B, refrigerant at the discharge pressure Pdh2, supplied from the port 15, is about to leak into the port 14 via the clearance between the piston 33 and the second body 12, the leakage of the refrigerant is sealed by the seal ring 41 pressurized for deformation by the discharge pressure Pdh2. Further, although refrigerant at the discharge pressure Pdh, supplied from the port 13, leaks into the guide 19 via a portion where the guide 19 and the first body 11 are in sliding contact with each other, and leaks further into the port 14 via the check valve 21, this slight leakage is negligible since the leakage does not adversely affect the operations of the control valve 10 and the compressor.

As described above, the piston 33 of the second control valve 10B senses the differential pressure A P across the first control valve 10A, which is generated by refrigerant passing through the first control valve 10A set to the predetermined passage cross-sectional area, and the second control valve 10B controls the flow rate of refrigerant supplied to the crankcase such that the differential pressure A P is held constant. This enables the control valve 10 to control the compressor such that refrigerant is discharged at a flow rate corresponding to control current supplied to the solenoid section 10C.

Next, operation for stopping the control valve 10 in the above control state will be described hereinafter.

When the supply of the control current to the solenoid section 10C is suddenly stopped so as to stop the compressor, the solenoid force which has set the lift amount of the first control valve 10A is lost, whereby the first control valve 10A is instantaneously fully closed by the urging force of the spring 20. This allows the first valve element 18 to pull the piston 33 downward, as viewed in FIG. 1, to thereby fully close the second control valve 10B forcibly. This causes the compressor to shift to the minimum displacement operation.

When the compressor shifts to the minimum displacement operation, the discharge pressure Pdh from the compressor quickly decreases, but the discharge pressure Pdl at the refrigerant outlet port of the compressor progressively decreases since the first valve element 18 is in the fully-closed state. Therefore, the discharge pressure Pdl sometimes becomes higher than the discharge pressure Pdh immediately after stoppage of the control current to the solenoid section 10C. In this case, the first control valve 10A is held in the fully-closed state since the discharge pressure Pdl acts on the first valve element 18 and the check valve 21 in the valve-closing directions, while the second control valve 10B is held in the fully-open state without being caused to operate by the discharge pressure Pdl since the piston 33 and the first valve element 18 engaged with the piston 33 have the same pressure-receiving area. At this time, leakage of refrigerant from the port 14 to the port 15 via the clearance between the second body 12 and the piston 33 is sealed by the seal ring 40.

As described hereinbefore, even when the discharge pressure Pdl on the refrigerant outlet port side of the compressor becomes equal to or higher than the discharge pressure Pdh on the discharge chamber side of the compressor, the first control valve 10A can be held in the fully-closed state, which enables the first control valve 10A to act similarly to the check valve conventionally provided in the refrigerant outlet port of the compressor, and the second control valve 10B can be held in the fully-open state, which enables the compressor to positively shift to the minimum displacement operation state. This means that the check valve conventionally provided in the refrigerant outlet port of the compressor is dispensed with.

FIG. 3 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a second embodiment. In FIG. 3, component elements which have functions identical to or equivalent to those of the component elements appearing in FIG. 1, are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 70 according to the second embodiment is distinguished from the control valve 10 according to the first embodiment in that it is not equipped with the check valve 21 formed on the guide 19.

More specifically, in a first control valve 70A of the control valve 70, the guide 19 connected to the first valve element 18 for operation in unison therewith comprises a hollow cylindrical portion having a closed end, and a sliding portion which is integrally formed with the hollow cylindrical portion, and radially outwardly extends from an open end of the hollow cylindrical portion such that an outer peripheral surface of the extended portion slides on the inner wall surface of the first body 11. The shaft 54 of a solenoid section 70C extends into the hollow cylindrical portion for abutment with the closed end of the hollow cylindrical portion. Further, the guide 19 has an intercommunicating hole 19 a formed in a side of the hollow cylindrical portion such that pressure in the solenoid section 70C always becomes equal to the discharge pressure Pdh on the upstream side of the first control valve 70A. The other features of configuration of the control valve 70 according to the second embodiment are the same as those of the control valve 10 according to the first embodiment.

In the control valve 70 configured as above, first, when the solenoid section 70C is not energized, as shown in FIG. 3, the first control valve 70A is fully closed by the urging force of the spring 20, and a second control valve 70B is fully open since the first valve element 18 of the first control valve 70A forcibly pulls the piston 33 that senses a differential pressure across the first control valve 70A, in the valve-opening direction. This allows the discharge chamber and the crankcase to communicate with each other via the second control valve 70B, and hence the compressor is in the minimum displacement operation state.

Then, when control current is supplied to the solenoid section 70C, as the control current increases, the shaft 54 of the solenoid section 70C lifts the first valve element 18 of the first control valve 70A, whereby the first control valve 70A starts to be opened, and the second control valve 70B starts to shift in the valve-closing direction in an interlocked manner. After that, when the second control valve 70B is closed, the compressor is placed in the maximum displacement operation state. After the second control valve 70B is closed, the first control valve 70A is positioned in a lift position corresponding to the control current, and is set to a passage cross-sectional area corresponding to the control current, without being interlocked with the second control valve 70B.

When the compressor shifts to the maximum displacement operation state, the flow rate of refrigerant passing through the first control valve 70A increases. When the differential pressure across the first control valve 70A becomes equal to or larger than a predetermined value, the piston 33 that senses the differential pressure acts to open the second control valve 70B to make the displacement of the compressor variable.

If the solenoid section 70C is deenergized when the control valve 70 is in the control state, the first control valve 70A is fully closed instantaneously by the spring 20, and the second control valve 70B is constrained and forcibly fully opened during transition of the first control valve 70A to the fully-closed state. This sharply decreases the discharge pressure Pdh in the discharge chamber while progressively decreasing the discharge pressure Pdl at the refrigerant outlet port of the compressor, to thereby invert the relationship between the discharge pressure Pdh and the discharge pressure Pdl. However, the first valve element 18 and the piston 33 integrally engaged with each other have the same pressure-receiving area, and are insensitive to the discharge pressure Pdl, so that even if the discharge pressure Pdl becomes higher than the discharge pressure Pdh, the fully-closed state of the first control valve 70A and the fully-open state of the second control valve 70B are maintained.

FIG. 4 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a third embodiment. In FIG. 4, component elements which have functions identical to or equivalent to those of the component elements appearing in FIGS. 1 and 3, are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 80 according to the third embodiment is distinguished from the control valves 10 and 70 according to the first and second embodiments in that its construction is simplified by forming the two ports 13 and 15 through which refrigerant from the discharge chamber is introduced, into a common port.

More specifically, in the control valve 80, a first control valve 80A thereof has the first valve element 18 disposed on the upstream side of the first valve seat 17. The first valve element 18 has a through hole axially formed therethrough, and a hollow cylindrical portion 81 extending via a valve hole is rigidly press-fitted into the through hole. The hollow cylindrical portion 81 is integrally formed with a piston 82 axially slidably disposed within the first body 11. A hollow part of the hollow cylindrical portion 81 extends into the piston 82 such that the hollow part communicates with a side opposite to the side where the hollow cylindrical portion 81 is formed, and the hollow part extending through the piston 82 forms a refrigerant passage 83 through which the discharge pressure Pdh is introduced into the solenoid section 80C. Further, the piston 82 is urged in the valve-closing direction of the first control valve 80A by the spring 20 disposed between the piston 82 and an end face of the second body 12. Furthermore, the piston 82 is configured such that it has a large outer diameter on the side toward a second control valve 80B, whereby the first control valve 80A has a valve structure in which when the urging force of the spring 20 exceeds the urging force of the solenoid section 80C, the clearance between the first body 11 and the piston 82 is sealed.

In the second control valve 80B, the port 16 is formed in the center of a foremost end of the second body 12, and an inner open end of the port 16 forms a second valve seat. The second valve element 32 is disposed in a manner movable to and away from the second valve seat. The second valve element 32 is integrally formed with a hollow cylindrical body 84 axially slidably disposed within the second body 12. The hollow cylindrical body 84 has the same outer diameter as that of the piston 82 of the first control valve 80A, and at the same time is formed with a plurality of intercommunicating holes. Further, the hollow cylindrical body 84 has the first valve seat 17 of the first control valve 80A rigidly press-fitted into the inside thereof. Disposed between the first valve seat 17 and the piston 82 is the spring 37 for urging the hollow cylindrical body 84 in the valve-closing direction of the second control valve 80B.

In the control valve 80 configured as above, first, when the solenoid section 80C is not energized, as shown in FIG. 4, the first control valve 80A is fully closed by the urging force of the spring 37 since the first valve seat 17 has the first valve element 18 brought into abutment therewith, while in the second control valve 80B, the second valve element 32 integrally formed with the hollow cylindrical body 84 is held in a fully-open position since the first valve seat 17 rigidly fixed to the hollow cylindrical body 84 has the first valve element 18 brought into abutment therewith by the spring 37. This allows the discharge chamber and the crankcase to communicate with each other via the second control valve 80B, and hence the compressor is in the minimum displacement operation state.

Then, when control current is supplied to the solenoid section 80C, as the control current increases, the shaft 54 of the solenoid section 80C pushes the piston 82 of the first control valve 80A upward, as viewed in FIG. 4, whereby the movable parts of the first control valve 80A and the second control valve 80B move in unison in the valve-closing direction of the second control valve 80B. After that, when the second valve element 32 is seated on the second valve seat to close the second control valve 80B, the compressor shifts to the maximum displacement operation state. Then, when the piston 82 is pushed in the valve-closing direction of the second control valve 80B, the first valve element 18 is progressively lifted from the first valve seat 17 to progressively open the first control valve 80A. Subsequently, the first valve element 18 stops at a lift position corresponding to the control current, and the first control valve 80A is set to a passage cross-sectional area corresponding to the control current.

When the compressor shifts to the maximum displacement operation state, the flow rate of refrigerant passing through the first control valve 80A increases to generate a differential pressure across the first control valve 80A. This differential pressure is received by the cross-sectional areas of the first valve seat 17 and the hollow cylindrical body 84 forming a differential pressure-sensing section. When the differential pressure becomes equal to or larger than a predetermined value, the first valve seat 17 and the hollow cylindrical body 84, which sense the differential pressure, act to open the second control valve 80B to make the displacement of the compressor variable.

Now, when the discharge pressure Pdh increases while the compressor is being controlled at a predetermined displacement, the second control valve 80B operates in the valve-opening direction to control the capacity in a direction of decreasing the same, whereas when the discharge pressure Pdh decreases, the second control valve 80B operates in the valve-closing direction to control the capacity in a direction of increasing the same. At this time, although the first control valve 80A as well changes in accordance with the change in the discharge pressure Pdh, the displacement of the compressor is determined depending on the control balance between the first control valve 80A and the second control valve 80B, and is substantially set to a predetermined displacement.

If the solenoid section 80C is suddenly deenergized when the control valve 80 is in the control state, the first control valve 80A is fully closed instantaneously by the spring 20, and the second control valve 80B is constrained and forcibly fully opened during transition of the first control valve 80A to the fully-closed state. This sharply decreases the discharge pressure Pdh in the discharge chamber while progressively decreasing the discharge pressure Pdl in the refrigerant outlet port of the compressor, so that the discharge pressure Pdl at the refrigerant outlet port presently becomes higher than the discharge pressure Pdh in the discharge chamber. However, the pressure-receiving area at which the first valve seat 17, the first valve element 18, and the hollow cylindrical body 84, which are integrally engaged with each other, receive the discharge pressure Pdl in the upward direction, as viewed in FIG. 4, is the same as the pressure-receiving area at which the piston 33 receives the discharge pressure Pdl in the downward direction, as viewed in FIG. 4, so that the control valve 80 has a structure insensitive to the discharge pressure Pdl. Accordingly, even if the discharge pressure Pdl becomes higher than the discharge pressure Pdh, the fully-closed state of the first control valve 80A and the fully-open state of the second control valve 80B are maintained.

FIG. 5 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a fourth embodiment. In FIG. 5, component elements which have functions identical to or equivalent to those of the component elements appearing in FIG. 4, are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 90 according to the fourth embodiment is distinguished from the control valve 80 according to the third embodiment in that the control valve 90 is improved in speed at which it returns again to the control state after suddenly transitioning from the control state to the stopped state.

More specifically, in the control valve 90, the first valve element 18 and the hollow cylindrical portion 81 are integrally formed with each other, and the refrigerant passage 83 is linearly axially formed therethrough. The hollow cylindrical portion 81 is fixed to the piston 82 in a manner extending therethrough, and an open end of the refrigerant passage 83 on the side toward a solenoid section 90C is opened and closed by an end face of the shaft 54. With this configuration, the control valve 90 has a valve structure in which when the solenoid section 90C is not energized as shown in FIG. 5, the port 13 through which the discharge pressure Pdh is introduced, and the inside of the solenoid section 90C are communicated with each other, whereas when the solenoid section 90C is energized, the communication between the port 13 through which the discharge pressure Pdh is introduced, and the inside of the solenoid section 90C is blocked.

In the construction described above, the basic operation of the control valve 90 is the same as that of the FIG. 4 control valve 80 according to the third embodiment. However, when the solenoid section 90C is energized immediately after the control valve 90 has suddenly shifted or transitioned from the control state to the stopped state, to return again to the control state, the shaft 54 lifts the first valve element 18 while closing the open end of the hollow cylindrical portion 81. When the first valve element 18 is lifted even in a slight degree, the piston 82 as well is lifted accordingly, to open the clearance between the piston 82 and the first body 11, and refrigerant at still high discharge pressure Pdl is introduced into a space formed between the piston 82 and the solenoid section 90C. This causes an urging force for causing a first control valve 90A to operate in the valve-opening direction to act on the piston 82, which helps the solenoid section 90C cause the first control valve 90A to operate in the valve-opening direction. This shortens a time period required for fully opening a second control valve 90B, and makes it possible for the control valve 90 to return to its original control state sooner.

FIG. 6 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a fifth embodiment. In FIG. 6, component elements which have functions identical to or equivalent to those of the component elements appearing in FIGS. 1 and 5, are designated by identical reference numerals, and detailed description thereof is omitted.

In the control valve 100 according to the fifth embodiment, the structure of the control valve 90 according to the fourth embodiment is applied to the control valve 10 according to the first embodiment, which has the ports 13 and 15 formed independently of each other for introducing refrigerant into the first control valve 10A and the second control valve 10B, respectively, and further the control valve 90 according to the fourth embodiment is improved in that the pressure Pc in the crankcase acts on the second control valve 90B in the valve-opening direction.

More specifically, in the control valve 100, the second valve seat 31 forming the port 15 for introducing the discharge pressure Pdh2 is formed on the foremost end of the second body 12, and the second valve element 32 is axially movably supported by the second body 12 in a manner opposed to the second valve seat 31. An end of the second valve element 32 on the side toward a solenoid section 100C extends up to a chamber of the port 13 into which the discharge pressure Pdh is introduced, and an engaging portion 101 held by a closing portion of the hollow cylindrical body 84 of the first control valve 100A is rigidly press-fitted into the end of the second valve element 32. Disposed between the engaging portion 101 and the first valve element 18 is the spring 35 for urging the second valve element 32 in the valve-closing direction. Further, an end face of the engaging portion 101 on a side opposite from a side receiving the spring 35 is tapered such that when the second valve element 32 is seated on the second valve seat 31, the clearance between the second valve element 32 and the second body 12 supporting the second valve element 32 can be closed.

In the control valve 100 configured as above, operation thereof is substantially the same as that of the control valve 90 according to the fourth embodiment. However, a second control valve 100B is configured such that the discharge pressure Pdh and the discharge pressure Pdh2, which are approximately equal to each other, are applied to axially opposite ends of the second valve element 32, respectively. This allows the control valve 100 to perform control operation in response to a differential pressure between the pressures that are applied to the first valve seat 17 and the hollow cylindrical body 84 from axially opposite sides, without being adversely affected by the pressure Pc from the crankcase.

FIG. 7 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a sixth embodiment. In FIG. 7, component elements which have functions identical to or equivalent to those of the component elements appearing in FIG. 1 are designated by identical reference numerals, and detailed description thereof is omitted.

In the control valve 110 according to the sixth embodiment, although its first control valve 110A and solenoid section 110C have quite the same constructions as the constructions of those of the control valve 10 according to the first embodiment, a second control valve 110B has a construction different from that of the second control valve 10B.

More specifically, in the second control valve 110B of the control valve 110, the differential pressure-sensing section is formed by a valve element-holding portion 111 that holds the second valve element 32, and a bellows 112 that has axially opposite ends thereof tightly connected to an upper end, as viewed in FIG. 7, of the valve element-holding portion 111 and an upper end, as viewed in FIG. 7, of the spring-receiving portion 36, respectively, such that the bellows 112 can axially extend and contract. In this differential pressure-sensing section as well, similarly to the piston 33 of the second control valve 10B, the valve element-holding portion 111 causes the second valve element 32 to axially move according to a differential pressure between the discharge pressure Pdh2 and the discharge pressure Pdl, whereby the valve lift of the second control valve 110B can be adjusted. Further, since the bellows 112 partitions between the port 15 at the discharge pressure Pdh2 and the port 14 at the discharge pressure Pdl, it is possible to completely prevent leakage of refrigerant from being caused by the differential pressure between the pressures Pdh2 and Pdl. This construction of the second control valve 110B makes it possible to dispense with the seal rings 40 and 41 used in the second control valve 10B.

In the control valve 110 constructed as above, operation thereof is the same as that of the control valve 10 according to the first embodiment, and detailed description thereof is omitted.

FIG. 8 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to a seventh embodiment. In FIG. 8, component elements which have functions identical to or equivalent to those of the component elements appearing in FIG. 3 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 120 according to the seventh embodiment is distinguished from the control valve 70 according to the second embodiment in that it has a simpler construction. More specifically, in a first control valve 120A of the control valve 120, respective both ends of the first valve element 18 and the guide 19 connected thereto are in sliding contact with the respective inner walls of the first body 11 and the second body 12 such that the first valve element 18 and the guide 19 can axially move in a stable state. Further, the guide 19 has a communication hole 121 formed through a portion close to a portion with which the shaft 54 of a solenoid section 120C is in abutment.

A second control valve 120B includes the piston 33 that senses the differential pressure between the discharge pressure Pdh2 at the port 15 and the discharge pressure Pdl at the port 14, and a shaft 122 fixed to the piston 33. One end of the shaft 122 forms the second valve element 32 of the second control valve 120B, while the other end thereof forms an engaging portion with which the first valve element 18 of the first control valve 120A is engaged when the solenoid section 120C is not energized, and cooperates with the first valve element 18 to form a valve element of a valve that opens and closes a refrigerant passage axially formed through the centers of the first valve element 18 and the guide 19. Further, the spring 37 urging the piston 33 in the valve-closing direction of the second control valve 120B is disposed between the seal ring 40 and the first valve element 18 of the first control valve 120A. Although the spring 37 should urge the piston 33 with respect to the second body 12, it is configured to urge the piston 33 with respect to the first valve element 18 so as to reduce a spring load and simplify the construction.

According to the control valve 120, when the solenoid section 120C is not energized, the spring 20 pushes the first valve element 18 downward, as viewed in FIG. 8, to fully close the first control valve 120A, and the first valve element 18 pulls the shaft 122 downward, as viewed in FIG. 8, to fully open the second control valve 120B. At this time, since the first valve element 18 and the engaging portion of the shaft 122 are tightly engaged with each other, the refrigerant passage formed through the centers of the first valve element 18 and the guide 19 is closed, and refrigerant at the discharge pressure Pdh leaks through where the guide 19 and the first body 11 are in sliding contact with each other, so that pressure in the solenoid section 120C is close to the discharge pressure Pdh.

When the solenoid section 120C is energized, the shaft 54 pushes the first valve element 18 upward, as viewed in FIG. 8, via the guide 19 to open the first control valve 120A. The piston 33 as well is pushed upward, as viewed in FIG. 8, by the spring 37 in a manner interlocked with the pushing of the first valve element 18 upward, and when the second valve element 32 is seated, the second control valve 120B is fully closed. When the first valve element 18 is further pushed upward, the shaft 122 tightly engaged with the first valve element 18 comes to be away from the first valve element 18, so that the refrigerant passage formed through the centers of the first valve element 18 and the guide 19 is opened, and the solenoid section 120C communicates with the port 14 to make the pressure therein equal to the discharge pressure Pdl.

When supply of the control current to the solenoid section 120C is suddenly stopped from the control state in which the first valve element 18 is lifted to a predetermined value by the control current supplied to the solenoid 120C, the first control valve 120A is fully closed, and the second control valve 120B is fully opened, while fully closing the refrigerant passage formed in the centers of the first valve element 18 and the guide 19. This causes the compressor to shift to the minimum displacement operation state, whereby the discharge pressures Pdh and Pdh2 on the discharge chamber side are sharply decreased, and the discharge pressure Pdl in the solenoid section 120C leaks into the port 13 through where the guide 19 and the first body 11 are in sliding contact with each other, to make the discharge pressure Pdl close to the discharge pressures Pdh sharply decreased. As a result, the discharge pressures Pdh and Pdh2, which have been sharply decreased to become approximately equal to each other, are applied to the axially opposite ends of the movable part of the guide 19, the first valve element 18, and the piston 33, which are made integral with each other, and therefore the fully-closed state of the first control valve 120A and the fully-open state of the second control valve 120B are maintained almost only by the load of the spring 20.

FIG. 9 is a central longitudinal cross-sectional view of the construction of a control valve for a variable displacement compressor, according to an eighth embodiment. FIGS. 10A to 10C are views useful in explaining characteristics of a diaphragm, in which FIG. 10A shows a state in which no differential pressure is applied to the diaphragm, FIG. 10B shows a state in which the diaphragm is displaced by a differential pressure, and FIG. 10C shows a state in which the differential pressure is applied to the displaced diaphragm in a direction opposite to the direction of the displacement. FIGS. 11A and 11B are explanatory views showing the construction of a differential pressure-sensing section of the control valve according to the eighth embodiment, in which FIG. 11A shows a case where pressure in the discharge chamber of the compressor is higher than pressure in the refrigerant outlet port of the compressor, and FIG. 11B shows a case where the pressure in the discharge chamber is lower than the pressure in the refrigerant outlet port. In FIG. 9 and FIGS. 11A and 11B, component elements which have functions identical to or equivalent to those of the component elements appearing in FIG. 8 are designated by identical reference numerals, and detailed description thereof is omitted.

The control valve 130 according to the eighth embodiment is constructed by lower-cost component parts in place of the high-cost bellows 112 used in the control valve 110 according to the sixth embodiment, and high-cost cut parts used in the control valve 120 according to the seventh embodiment.

More specifically, in the control valve 130, almost all the component elements of a first control valve 130A and a second control valve 130B are pressed parts formed by pressing pipes, and the pressed parts are assembled by press-fitting or swaging.

In the first control valve 130A, a first body 131 having an end thereof inwardly bent is fixed to a solenoid section 130C by swaging the foremost end of the core 51 protruding from the yoke 59, and a second body 132 and a third body 133 having one end forming the first valve seat 17 are rigidly press-fitted into the first body 131. The third body 133 has the guide 19, which has a shape of a bell, axially slidably disposed therein, and a bell-shaped shaft-receiving portion 134 having the communication hole 121 is press-fitted into the guide 19. The guide 19 has the first valve element 18 externally fitted on one end thereof on the side toward the second control valve 130B, and is urged in the valve-closing direction by the spring 20 disposed between the other end thereof and a protrusion formed on the inside of the body 133.

The second body 132 has an open end provided with a diaphragm 135 made of polyimide for sealing between the port 15 at the discharge pressure Pdh2 and the port 14 at the discharge pressure Pdl, and sensing the differential pressure between the discharge pressure Pdh2 and the discharge pressure Pdl. The diaphragm 135 has an outer peripheral portion thereof sandwiched between a first ring 136 and a second ring 137, and a central portion thereof sandwiched between a center disk 138 and a flange portion 139. The first ring 136 and the second ring 137 are fixed to the second body 132 together with a fourth body 140 by swaging open ends of the second body 132, in a state sandwiching the diaphragm 135 therebetween. On the other hand, the center disk 138 and the flange portion 139 are fixed to each other in a state sandwiching the diaphragm 135 therebetween, by press-fitting a shaft 141 into a central portion of the center disk 138 and the hollow cylindrical second valve element 32 of the second control valve 130B, integrally formed with the flange portion 139. It should be noted that portions of the first and second rings 136 and 137, sandwiching the outer peripheral portion and the central portion of the diaphragm 135, are configured such that the inner diameter of the first ring 136 is set to be larger than the inner diameter of the second ring 137, and the outer diameter of the center disk 138 is set to be larger than the outer diameter of the flange portion 139. Further, the first ring 136 has a stepped portion, and a portion thereof extending from the stepped portion forms a stopper 142 for restricting the displacement of the diaphragm 135.

A cup-shaped fifth body 143 is press-fitted into the fourth body 140. The fifth body 143 has a valve hole of the second control valve 130B formed in the center of its bottom, and an opening of the valve hole forms the port 16 leading to the crankcase. The shaft 141 extends through the valve hole of the second control valve 130B, and a spring-receiving portion 144 is externally fitted on a foremost end of the shaft 141. Interposed between the bottom of the fifth body 143 and the spring-receiving portion 144 is the spring 37 for urging the second valve element 32 in the valve-closing direction.

According to the control valve 130 described above, when the solenoid section 130C is not energized, the spring 20 pushes the guide 19 downward, as viewed in FIG. 9, to fully close the first control valve 130A, and at the same time the guide 19 pulls the shaft 141 downward, as viewed in FIG. 9, until the center disk 138 is brought into abutment with the stopper 142, to fully open the second control valve 130B. At this time, since the guide 19 and the shaft 141 are tightly engaged with each other, and refrigerant at the discharge pressure Pdh leaks into the first body 131 through where the guide 19 and the third body 133 are in sliding contact with each other, so that pressure in the solenoid section 130C is close to the discharge pressure Pdh.

When the solenoid section 130C is energized, the shaft 54 pushes the first valve element 18 upward, as viewed in FIG. 9 via the guide 19, to open the first control valve 130A. The shaft 141 as well is pulled upward, as viewed in FIG. 9, by the spring 37 in a manner interlocked with the pushing of the first valve element 18 upward, and when the second valve element 32 is seated, the second control valve 130B is fully closed. When the first valve element 18 is further pushed upward, the shaft 141 tightly engaged with the guide 19 comes to be away from the guide 19, so that the solenoid section 130C communicates with the port 14 via a hole in the center of the guide 19, through which the shaft 141 extends, and the communication hole 121 of the shaft-receiving portion 134, whereby pressure in the solenoid section 130C becomes equal to the discharge pressure Pdl.

After that, when control current is held at a predetermined value, the first valve element 18 is stopped at a position where the urging force of the solenoid section 130C corresponding to the predetermined value and the urging force of the spring 20 against the solenoid force are balanced. The first valve element 18 is lifted from the first valve seat 17 and is stopped, whereby the first control valve 130A is set as to a refrigerant passage thereof such that the refrigerant passage has a predetermined passage cross-sectional area, so that refrigerant at the discharge pressure Pdh, introduced into the port 13, is allowed to flow through the refrigerant passage having the predetermined passage cross-sectional area, and refrigerant at the discharge pressure Pdl is discharged from the port 14 into the refrigerant outlet port of the compressor. When refrigerant flows through the first control valve 130A, a predetermined differential pressure ΔP is generated across the first control valve 130A. Since the discharge pressure Pdh2 is approximately equal to the discharge pressure Pdh, the differential pressure ΔP generated across the first control valve 130A is sensed by the diaphragm 135. The diaphragm 135 drives the second valve element 32 that moves in unison therewith, whereby it controls the flow rate of refrigerant supplied via the second control valve 130B to the crankcase. Thus, the control valve 130 provides control such that the compressor discharges refrigerant at a flow rate corresponding to the control current supplied to the solenoid section 130C.

When the supply of the control current to the solenoid section 130C is suddenly stopped from the control state in which the first valve element 18 is lifted to a predetermined value by the control current supplied to the solenoid 130C, the first control valve 130A is fully closed, and the second control valve 130B is fully opened, while closing the hole in the center of the guide 19. This causes the compressor to shift to the minimum displacement operation state, whereby the discharge pressures Pdh and Pdh2 on the discharge chamber side are sharply decreased, and the discharge pressure Pdl in the solenoid section 130C leaks into the port 13 through the guide 19 and the first body 11 are in sliding contact with each other, to make the discharge pressure Pdl close to the discharge pressures Pdh sharply decreased. As a result, the discharge pressure Pdl progressively decreased is applied to the guide 19 from the side toward the port 14, whereas the discharge pressures Pdh sharply decreased is applied to the inside of the guide 19, and hence the differential pressure therebetween and the load of the spring 20 maintains the fully-closed state of the first control valve 130A and the fully-open state of the second control valve 130B.

Now, when attention is paid to the differential pressure-sensing section, it is known that an effective pressure-receiving area of the diaphragm 135 is changed according to the stroke of the displacement thereof. As shown in FIG. 10A, the effective pressure-receiving area of the diaphragm 135 depends on the area of a circle a diameter (effective diameter b) of which is the distance between the centers of curvature circles a of respective corrugated portions. Here, when pressure P1 applied from above, as viewed in the figures, becomes larger than pressure P2 applied from below, as viewed in the same, a central portion of the diaphragm 135 is displaced downward, as viewed in FIG. 10B. At this time, since an inner peripheral portion of each corrugated portion is also displaced together with the central portion, the curvature of the corrugated portion is increased, and the center of the curvature moves inward, whereby the effective diameter becomes an effective diameter b1 smaller than the effective diameter b, to decrease the effective pressure-receiving area. Here, as shown in FIG. 10C, when the pressure P2 becomes larger than the pressure P1 in the state in which the central portion of the diaphragm 135 is displaced, the corrugated portion alone expands to swell toward the pressure P1, which causes the center of the curvature to move outward such that the effective diameter becomes an effective diameter b2 larger than the effective diameter b, to increase the effective pressure-receiving area.

This situation corresponds to a case where the supply of the control current to the solenoid section 130C is suddenly stopped to make the discharge pressure Pdl on the downstream side of the first control valve 130A higher than the discharge pressure Pdh2 on the upstream side thereof. In such a case, the second valve element 32 that moves in unison with the diaphragm 135 is caused to act in the valve-closing direction by the differential pressure between the discharge pressure Pdl and the discharge pressure Pdh2. More specifically, when the solenoid section 130C is deenergized, the second control valve 130B is fully opened by the load of the spring 20, but immediately after that, when the discharge pressure Pdl becomes higher than the discharge pressure Pdh2, the diaphragm 135 responsive to the differential pressure acts on the second control valve 130B in the valve-closing direction. Therefore, particularly when the differential pressure is large, it is impossible to maintain the fully-closed state of the first control valve 130A and the fully-open state of the second control valve 130B.

In contrast, the control valve 130 is configured such that when the discharge pressure Pdl becomes higher than the discharge pressure Pdh2, the force that acts on the second control valve 130B in the valve-closing direction is inhibited from being increased. To this end, it is only required that an effective diameter c2 of the diaphragm 135 obtained when the discharge pressure Pdl is higher than the discharge pressure Pdh2 is made smaller than an effective diameter c1 of the diaphragm 135 obtained when the discharge pressure Pdh2 is higher than the discharge pressure Pdl. This is realized, as shown in FIGS. 11A and 11B in an enlarged form, by making the respective inner diameters of the first and second rings 136 and 137 which sandwich the respective surfaces of the diaphragm 135, different from each other, and making the respective outer diameters of the center disk 138 and the flange portion 139 which also sandwich the respective surfaces of the diaphragm 135, different from each other. More specifically, the inner diameter of the stepped portion of the first ring 136 is set to be larger than that of the second ring 137, and at the same time the outer diameter of the center disk 138 is set to be larger than that of the flange portion 139. However, the effective diameter c1 is set to be equal to the inner diameter of the first valve seat of the first control valve 130A such that when the first control valve 130A is in the closed state, the diaphragm 135 has the same pressure-receiving area as the pressure-receiving area of the first valve element 18 for receiving the discharge pressure Pdl. It should be noted that the distance between the inner periphery of the first ring 136 and the outer periphery of the center disk 138 is set to be equal to the distance between the inner periphery of the second ring 137 and the outer periphery of the flange portion 139.

As a result, when Pdl<Pdh2 holds, as shown in FIG. 11A, the corrugated portion of the diaphragm 135 is defined by the center disk 138 having a larger outer diameter and the first ring 136 having a larger inner diameter, and the effective pressure-receiving area of the diaphragm 135 at this time is determined by the effective diameter c1. On the other hand, when Pdl>Pdh2 holds, as shown in FIG. 11B, the corrugated portion of the diaphragm 135 is defined by the flange portion 139 having a smaller outer diameter and the second ring 137 having a smaller inner diameter, and the effective pressure-receiving area of the diaphragm 135 at this time is determined by the effective diameter c2. As described above, when deenergization of the solenoid section 130C causes transition from the state in which the discharge pressure Pdl is lower than the discharge pressure Pdh2 to the state in which the discharge pressure Pdl is higher than the discharge pressure Pdh2, the effective pressure-receiving area of the diaphragm 135 is changed to be made smaller, whereby it is possible to reduce the force caused to act on the second control valve 130B in the valve-closing direction by the differential pressure, which makes it possible to smoothly perform a stopping operation of the automotive air conditioner.

Since the control valve for a variable displacement compressor, according to the present invention, is configured such that it is insensitive to pressure on the downstream side of the first control valve, pressure at the refrigerant outlet port is prevented from acting on the second control valve in the direction of increasing the displacement of the compressor even when the pressure at the refrigerant outlet port becomes higher than the pressure in the discharge chamber. This makes it possible to dispense with the check valve conventionally provided at the refrigerant outlet port of the compressor, which is advantageous in that it is possible to reduce the costs of the compressor.

The foregoing is considered as illustrative only of the principles of the present invention. Further, since numerous modifications and changes will readily occur to those skilled in the art, it is not desired to limit the invention to the exact construction and applications shown and described, and accordingly, all suitable modification and equivalents may be regarded as falling within the scope of the invention in the appended claims and their equivalents. 

1. A control valve for a variable displacement compressor, including a first control valve that controls a flow rate of refrigerant which said first control valve allows to flow from a discharge chamber of the compressor to a refrigerant outlet port of the compressor, a second control valve that controls a flow rate of refrigerant which said second control valve allows to flow from the discharge chamber into a crankcase of the compressor based on a differential pressure across said first control valve, to change a displacement of the compressor, to thereby control the flow rate of the refrigerant allowed to flow by said first control valve to be constant, and a solenoid section that sets the flow rate of refrigerant which said first control valve is to allow to flow, wherein the control valve is made insensitive to pressure on a downstream side of said first control valve, and when said solenoid section is in a non-energized state, said first control valve is in a fully closed state, and said second control valve is in a fully open state, said second control valve is forcibly held in the fully open state even when the pressure on the downstream side of said first control valve is equal to or higher than pressure on an upstream side of said first control valve.
 2. The control valve according to claim 1, wherein: said first control valve has a first valve element that has a lift position thereof set according to an urging force of said solenoid section and is urged in a valve-closing direction against the urging force of said solenoid section; said second control valve has a differential pressure-sensing section that has a same pressure-receiving area as a pressure-receiving area of part of said first valve element, which receives the pressure on the downstream side of said first control valve when said first valve element is in a closed position, and senses a differential pressure between the pressure on the upstream side of said first control valve and the pressure on the downstream side of said first control valve, which are generated across said first control valve, and a second valve element that is held by said differential pressure-sensing section; and upon transition of said solenoid section from a control state in which said solenoid section sets said first control valve to a non-energized state, said first valve element being shifted to a fully closed position by the urging force engages with said differential pressure-sensing section which has been away from said first valve element during the control state, to cause said second valve element to shift to a fully open position.
 3. The control valve according to claim 2, wherein said second control valve, said first control valve, and said solenoid section are arranged along a same axis in order, wherein said first control valve has a first port formed on a side toward said solenoid section, for introducing refrigerant from the discharge chamber, a second port formed on a side toward said second control valve, for discharging refrigerant into the refrigerant outlet port, a first valve seat provided between said first port and said second port, and said first valve element disposed on a downstream side of said first valve seat in a manner movable to and away from said first valve seat, and said second control valve has a third port separated from said second port by said differential pressure-sensing section, for introducing refrigerant from the discharge chamber, a fourth port formed on a side opposite from said first control valve and along an axis thereof, for discharging refrigerant into the crankcase, said differential pressure-sensing section urged in a valve-closing direction, a second valve seat disposed in said fourth port, and said second valve element disposed on an upstream side of said second valve seat and held by said differential pressure-sensing section in a manner movable to and away from said second valve seat.
 4. The control valve according to claim 3, wherein said first control valve has a guide slidably disposed along an axis of said first valve element within a space communicating with said first port and connected to said first valve element via a valve hole, for guiding an axial motion of said first valve element, and a spring disposed between said guide and said first valve seat, for urging said first valve element in the valve-closing direction.
 5. The control valve according to claim 4, wherein said guide has a same pressure-receiving area as a pressure-receiving area of part of said first valve element, which receives the pressure on the upstream side of said first control valve when said first valve element is in the closed position, and receives the pressure on the upstream side of said first control valve in a valve-closing direction of said first control valve.
 6. The control valve according to claim 5, wherein said first valve element is slidably disposed along an axis of said first control valve within a space communicating with said second port.
 7. The control valve according to claim 5, wherein said first valve element and said guide connected thereto have a refrigerant passage axially formed therethrough, and a check valve for closing said refrigerant passage when pressure in said refrigerant passage on a side toward said second port has become higher than pressure in said refrigerant passage on a side toward said solenoid section.
 8. The control valve according to claim 6, wherein said first valve element and said guide connected thereto have a refrigerant passage axially formed therethrough, and a valve that engages with said differential pressure-sensing section to thereby close said refrigerant passage when said first valve element forcibly shifts said second valve element to a fully open position.
 9. The control valve according to claim 4, wherein said guide is provided with an intercommunicating hole for making pressure in said solenoid section equal to the pressure on the upstream side of said first control valve.
 10. The control valve according to claim 3, wherein said second control valve has a piston as said differential pressure-sensing section for receiving pressure from said second port and pressure from said third port at axially opposite ends thereof to operate according to a differential pressure between the pressures, and a spring urging said piston in the valve closing direction.
 11. The control valve according to claim 10, wherein said spring is disposed between said piston and said first valve element.
 12. The control valve according to claim 10, wherein said second control valve has a film-like seal ring which is disposed on at least one of open ends of a clearance where the clearance formed between said piston and a body that axially movably holds said piston opens toward said second port and said third port, for sealing the clearance by the pressure from said second port or said third port.
 13. The control valve according to claim 12, wherein said second control valve has a valve element base portion-accommodating portion formed in an end face of said piston, opposed to said second valve seat, and a base portion of said second valve element is accommodated in said valve element base portion-accommodating portion in a state urged in the valve-closing direction and in a manner prevented from coming off.
 14. The control valve according to claim 3, wherein said second control valve has a bellows that has axially opposite ends thereof tightly connected to said differential pressure-sensing section holding said second valve element, and a body that axially movably accommodates said differential pressure-sensing section, said bellows being capable of axially extending and contracting while sealing said third port from said second port.
 15. The control valve according to claim 3, wherein said second control valve has a diaphragm as said differential pressure-sensing section, which is disposed between said second port and said third port in a manner sealing said second port from said third port, for receiving pressure from said second port and pressure from said third port at axially opposite surfaces thereof to cause said second valve element to operate by a differential pressure between the pressures.
 16. The control valve according to claim 15, wherein said diaphragm is tightly connected to a body in a state in which an outer periphery thereof is sandwiched between a first ring and a second ring, a central portion thereof being sandwiched between a center disk and a flange portion integrally formed with said second valve element having a hollow cylindrical shape, said diaphragm being fixed to a shaft fitted thereon in a manner axially extending therethrough, together with said center disk and said flange portion.
 17. The control valve according to claim 16, wherein said diaphragm is configured such that a pressure-receiving area thereof for receiving pressure when the pressure in said third port is higher than the pressure in said second port is equal to a pressure-receiving area of said first valve element for receiving the pressure from said second port when said first control valve is in a closed state.
 18. The control valve according to claim 17, wherein said diaphragm is configured such that an inner diameter of said second ring is made smaller than an inner diameter of said first ring, and an outer diameter of said flange portion is made smaller than an outer diameter of said center disk, whereby a pressure-receiving area of said diaphragm for receiving pressure when the pressure in said second port is higher than the pressure in said third port is made smaller than the pressure-receiving area for receiving pressure when the pressure from said third port is higher than the pressure from said second port.
 19. The control valve according to claim 2, wherein said second control valve, said first control valve, and said solenoid section are arranged along a same axis in order, and wherein said first control valve has a first port formed on a side toward said solenoid section, for discharging refrigerant into the refrigerant outlet port, a second port formed on a side toward said second control valve, for introducing refrigerant from the discharge chamber, a first valve seat disposed between said first port and said second port, and said first valve element disposed on an upstream side of said first valve seat in a manner movable to and away from said first valve seat, and wherein said second control valve has a third port formed on a side opposite from said first control valve and along an axis thereof, for discharging refrigerant introduced into said second port into the crankcase, a hollow cylindrical body that has said first valve seat fixed to an inside thereof and is axially movably disposed in a state urged in the valve-closing direction, forming said differential pressure-sensing section, and a second valve element integrally formed with said hollow cylindrical body, for opening and closing said third port.
 20. The control valve according to claim 19, wherein said first control valve includes a piston slidably disposed along an axis of said first valve element within a space communicating with said first port, and connected to said first valve element via a valve hole, for guiding an axial motion of said first valve element, and a spring disposed between said piston and a body accommodating said piston, for urging said first valve element in a valve-closing direction.
 21. The control valve according to claim 20, wherein said piston is configured to have a same outer diameter as that of said hollow cylindrical body of said second control valve, to thereby inhibit said piston from sensing the pressure on the downstream side of said first control valve when said first valve element is in the closed position.
 22. The control valve according to claim 20, wherein said first valve element and said piston connected thereto have a refrigerant passage axially formed therethrough for causing pressure of the refrigerant introduced into said second port to be received by an end face toward said solenoid section via said refrigerant passage.
 23. The control valve according to claim 22, wherein when a shaft of said solenoid section urges said first valve element, said refrigerant passage is closed by said shaft.
 24. The control valve according to claim 2, wherein said second control valve, said first control valve, and said solenoid section are arranged along a same axis in order, and wherein said first control valve has a first port formed on a side toward said solenoid section, for discharging refrigerant into the refrigerant outlet port, a second port formed on a side toward said second control valve, for introducing refrigerant from the discharge chamber, a first valve seat disposed between said first port and said second port, and said first valve element disposed on an upstream side of said first valve seat in a manner movable to and away from said first valve seat, and wherein said second control valve has a third port for discharging refrigerant into the crankcase, a fourth port formed on a side opposite from said first control valve and along an axis thereof, for introducing refrigerant from the discharge chamber, a hollow cylindrical body disposed within a space communicating with said second port and having said first valve seat rigidly fixed to an inside thereof in a state urged in the valve-closing direction, forming said differential pressure-sensing section, a second valve seat provided in said fourth port, and a second valve element that is disposed such that one end thereof is opposed to said second valve seat, and the other end thereof is urged in the valve closing direction with respect to said first valve element, for thereby being engaged with said hollow cylindrical body, and is movable to and away from said second valve seat.
 25. The control valve according to claim 24, wherein said first control valve includes a piston slidably disposed along an axis of said first valve element within a space communicating with said first port, and connected to said first valve element via a valve hole, for guiding an axial motion of said first valve element, and a spring disposed between said piston and a body accommodating said piston, for urging said first valve element in the valve-closing direction, and wherein said first valve element has a refrigerant passage axially formed therethrough which is closed by a shaft of said solenoid section when said shaft urges said first valve element.
 26. The control valve according to claim 25, wherein said piston is configured such that a portion thereof receiving the pressure at said first port in the valve-closing direction has a same outer diameter as that of said hollow cylindrical body of said second control valve, to thereby inhibit said piston from sensing the pressure on the downstream side of said first control valve when said first valve element is in the closed position. 